Direct torque path differential having spiderless pinions

ABSTRACT

A side pinion is disclosed for use with a differential of a mobile machine. The side pinion may have a body with a flat bottom and a flat top located at an end opposite the flat bottom. The side pinion may also have a plurality of gear teeth formed adjacent the flat top, and an arcuate outer surface connecting the plurality of gear teeth to the flat bottom.

TECHNICAL FIELD

The present disclosure relates generally to a differential and, moreparticularly, to a direct torque path differential having spiderlesspinions.

BACKGROUND

Machines, such as wheel loaders and haul trucks, generally include adrivetrain that provides power to traction devices of the machines. Thedrivetrain is made up of at least three different elements, including apower source (e.g., an engine), a transmission driven by the powersource, and a differential that divides power from the transmissionbetween paired traction devices. The differential allows the pairedtraction devices to be driven at different speeds to accommodate turningof the machine.

A differential generally consists of a main pinion gear that is drivenby the transmission to rotate a crown gear. A carrier housing is fixedto rotate with the crown gear, and includes two or more (e.g., four)different spider pinions internally located around a circumference ofthe carrier housing. The spider pinions are oriented radially inward androtatably disposed on a spider shaft (e.g., a cross having four shaftends), whose ends are connected to the carrier housing. Thus, when thecarrier housing rotates about its own axis, the spider pinions alsorotate about the same axis. In addition, each spider pinion spins aboutits own axis, which is oriented generally orthogonal to and passesthrough the axis of the carrier housing. A side gear is mounted at eachend of the carrier housing and intermeshes with the spider pinions. Theside gears are rotatable about the axis of the carrier housing, andconnected to half-shafts that extend outward from the differential torespective ones of the paired traction devices. With this configuration,an input rotation provided via the main pinion gear results in separaterotations of the traction devices with substantially equal torque.During straight travel of the machine over good ground conditions, bothtraction devices are driven at the same speed. During turning or poorground conditions, one traction device (e.g., the outside tractiondevice during a turn or the slipping traction device) speeds up as theremaining traction device slows down.

While acceptable for some applications, the conventional differentialcan be problematic in other applications. In particular, because of theconfiguration of typical spider pinions, a moment is created when teethof the spider pinions engage corresponding teeth of the side gears. Thismoment causes the spider pinions to tilt about the spider shaft ends.This tilting can restrict lubrication flow along the spider shaft (i.e.,inside bores of the spider pinions), and even cause mechanicalengagement between bore walls of the spider pinions and the spider shaftin some situations. The restricted lubrication and mechanical engagementcauses premature wear of the differential.

One attempt to extend a useful life of a differential is disclosed in USPatent Publication No. 2010/0151983 (the '983 publication) of Ziech etal. that published on Jun. 17, 2010. Specifically, the '983 publicationdiscloses a differential having a case, with a ring gear connected to anouter surface of the case and intermeshed with a pinion gear. Fourrecesses are formed within an inner surface of the case and directedradially inward into a hollow cavity in the case. The recesses areequally spaced around a circumference of the case. A wear cup is locatedwithin each of the recesses, and a tab of the wear cup is receivedwithin a slot in the case to prevent rotation of the wear cup. The wearcup has a flat base, and side walls that are perpendicular to the base.A side pinion is located within each of the wear cups. Each pinion has aflat heel end that fits inside the corresponding wear cup, a toe end,and a cylindrical wall that extends from the heel end to the toe end.The heel end directly contacts the flat base of the cup, and the sidewalls of the cup engage the cylindrical wall of the pinion to driverotation of the pinion about an axis of the case. Side gears are alsolocated within the case and mesh with the side pinions. The side gearsare hollow and include splines that mesh with corresponding splines ofhalf-shafts that protrude from the case. This design eliminates the needfor a spider shaft.

Although the differential of the '983 publication may not suffer frommechanical engagement between the side pinions and a spider shaft(because the differential of the '983 publication does not include aspider shaft), the differential may still be less than optimal. Inparticular, the moment discussed above that can be created by engagementof the side pinions with the side gears may still exist. This moment maycause the heel end of the side pinions to tilt within the cups, makingmechanical engagement between the cylinder wall of the side pinions andthe cup side walls possible. In the same manner described above, thisengagement may restrict lubrication of the toe end of the pinion gearand result in premature wear.

The disclosed differential is directed to overcoming one or more of theproblems set forth above and/or other problems of the prior art.

SUMMARY

In one aspect, the present disclosure is directed to a side pinion for adifferential. The side pinion may include a body with a flat bottom anda flat top located at an end opposite the flat bottom. The side pinionmay also include a plurality of gear teeth formed adjacent the flat top,and an arcuate outer surface connecting the plurality of gear teeth tothe flat bottom.

In another aspect, the present disclosure is directed to a differential.The differential may include an input gear, and a carrier fixedlyconnected to the input gear. The carrier may be configured to rotatetogether with the input gear about a primary axis, and may have aninternal annular surface with a plurality of cups formed therein thatare equally spaced apart around a circumference of the carrier. Thedifferential may also include a first side gear disposed inside thecarrier and configured to rotate about the primary axis, a second sidegear disposed in the carrier at an end opposite the first side gear andalso configured to rotate about the primary axis, and a side piniondisposed within each of the plurality of cups and intermeshed with bothof the first and second side gears. The side pinion may have a pluralityof gear teeth that protrude from an associated one of the plurality ofcups radially inward toward the primary axis, and an arcuate outersurface connected to the plurality of gear teeth at an axial transitionregion. The arcuate outer surface may conform to an inner contour of theplurality of cups.

In another aspect, the present disclosure is directed to a drivetrainfor a mobile machine having first and second traction devices located atopposing sides. The drivetrain may include a power source, atransmission driven by the power source, and a main pinion operativelyconnected to an output of the transmission. The drivetrain may alsoinclude a first half-shaft connected to the first traction device, asecond half-shaft connected to the second traction device, and adifferential driven by the main pinion to rotate the first and secondhalf-shafts with substantially equal torque. The differential may havean input gear intermeshed with the main pinion, and a carrier fixedlyconnected to the input gear and configured to rotate together with theinput gear about a primary axis. The carrier may have an internalannular surface with a plurality of cups formed therein that are equallyspaced apart around a circumference of the carrier. The differential mayalso have a first side gear disposed inside the carrier and havingexternal teeth at an inner end and an outer end connected to the firsthalf-shaft, a second side gear disposed inside the carrier and havingexternal teeth at an inner end and an outer end connected to the secondhalf-shaft, and a side pinion disposed within each of the plurality ofcups and intermeshed with the external teeth of both of the first andsecond side gears. The side pinion may have a plurality of gear teeththat protrude from an associated one of the plurality of cups radiallyinward toward the primary axis, and an arcuate outer surface connectedto the plurality of gear teeth at an axial transition region andconforming to an inner contour of the plurality of cups. The arcuateouter surface of the side pinion may be formed by rotating a polynomialcurve of third order or higher around an axis of the side pinion. Whenthe first and second side gears engage the plurality of gear teeth atopposing sides of the side pinion, a reaction force may be created, andthe polynomial curve is selected such that a gradient of the arcuateouter surface at an application point of the reaction force may be aboutaligned with the reaction force.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is an isometric illustration of an exemplary disclosed machine;

FIG. 2 is an isometric illustration of an exemplary disclosed drivetrainthat may be used with the machine of FIG. 1;

FIG. 3 is a cross-sectional illustration of an exemplary discloseddifferential that may be used with the drivetrain of FIG. 2;

FIGS. 4, 5, 6A, and 6B are plan, isometric view, cross-sectional, andenlarged illustrations, respectively, of an exemplary disclosed sidepinion that may form a portion of the differential of FIG. 3;

FIGS. 7 and 8 are cross-sectional illustrations of exemplary lubricationflows through the differential of FIG. 3;

FIGS. 9 and 10 are cross-sectional side and end view illustrations,respectively, of another exemplary disclosed differential that may beused with the drivetrain of FIG. 2;

FIGS. 11 and 12 are isometric and cross-sectional side viewillustrations, respectively, of an exemplary disclosed nested carrierthat may be used with the differential of FIGS. 8 and 9;

FIG. 13 is a cross-sectional illustration of another exemplary discloseddifferential that may be used with the drivetrain of FIG. 2;

FIG. 14 is an isometric side view illustration of another exemplarydisclosed nested carrier that may be used with the differential of FIG.13; and

FIG. 15 is a cross-sectional illustration of another exemplary discloseddifferential that may be used with the drivetrain of FIG. 2.

DETAILED DESCRIPTION

FIG. 1 illustrates an exemplary mobile machine 10. In the depictedembodiment, machine 10 is a wheel loader. It is contemplated, however,that machine 10 may embody another type of mobile machine such as anarticulated haul truck, an off-highway mining truck, a motor grader, oranother machine known in the art. Machine 10 may include an operatorstation 12, one or more traction devices 14 located at opposing sides ofmachine 10 that support operation station 12, and a drivetrain 16operatively connected to propel traction devices 14 in response to inputreceived via operator station 12.

As shown in FIG. 2, drivetrain 16 may be an assembly of components thattransfers power from a power source (e.g. an engine) 18 to tractiondevices 14. In the disclosed embodiment, these components include atransmission 20 that is operatively connected to and driven by powersource 18, one or more differentials 22 operatively connected betweenpairs of opposing traction devices 14, and one or more output shafts 24that connect transmission 20 to differentials 22. Transmission 20 may beconfigured to selectively vary a speed-to-torque ratio of the outputfrom power source 18 that is delivered to differential(s) 22. Eachdifferential 22 may be configured to provide a substantially equaltorque to the associated traction devices 14, allowing paired tractiondevices 14 to rotate at different speeds during turning of machine 10.

FIG. 3 illustrates an exemplary embodiment of differential 22. As can beseen in this figure, differential 22 is configured to receive input fromshaft 24 via a main pinion 26, and to direct outputs to traction devices14 (referring to FIG. 2) via first and second half-shafts 28 and 30. Inthe disclosed embodiment, differential 22 may have a primary axis 32,along which first and second half-shafts 28, 30 are aligned. Primaryaxis 32, in this embodiment, is oriented generally orthogonal to an axis34 of main pinion 26. It is contemplated, however, that in otherembodiments, primary axis 32 may be generally parallel with axis 34, ifdesired.

Main pinion 26 may be any type of gear known in the art that isconnectable to an end of shaft 24. Main pinion 26 is depicted in FIG. 3as a hypoid gear having a frustoconical shape and being welded to an endof shaft 24. Teeth formed within an exterior surface of main pinion 26have a spiraling trajectory, and axis 34 of main pinion 26 does not passthrough primary axis 32 of differential 22. It is contemplated that mainpinion 26 could alternatively embody a spiral bevel gear, a straightbevel gear, a single- or double-helical gear, or a spur gear, ifdesired.

First and second half-shafts 28, 30 may each have an inner endconfigured to connect with differential 22, and an outer end configuredto directly or indirectly connect with the associated traction device14. In some embodiments, the outer end of first and second half-shafts28, 30 connect directly with an intermediate speed reducer (e.g., afinal drive—not shown), which then connects with traction device 14.Other configurations may also be possible. The inner end of first andsecond half-shafts 28, 30 are shown as having external splines thatfacilitate connection to differential 22. In other embodiments, however,first and second half-shafts 28, 30 may additionally or alternativelyhave bolted flanges, keyways, and/or other means of connection.

Differential 22 may be an assembly of components that cooperate todivide the torque received from main pinion 26 between first and secondhalf-shafts 28, 30. The components may include, among other things, aninput gear 36 configured to mesh with main pinion 26; a carrier 38fixedly connected to rotate with input gear 36 about primary axis 32;first and second side gears 40, 42 connected to first and secondhalf-shafts 28, 30, respectively; and a plurality of side pinions 44that rotate with carrier 38 about axis 32 and simultaneously intermeshwith both of first and second side gears 40, 42. The rotational powerreceived from shaft 24 may pass through main pinion 26 to carrier 38 viainput gear 36. Carrier 38 may then transmit the rotational power throughside pinions 44 and first and second side gears 40, 42 to first andsecond half-shafts 28, 30, respectively.

Input gear 36, like main pinion 26, may also be a hypoid gear havingteeth that intermesh with the teeth of main pinion 26. As shown in FIG.3, the teeth of input gear 36 are located at an axial end adjacentcarrier 38. In other embodiments, however, the teeth of input gear 36could alternatively be located at an opposing axial end or on an outerannular edge, if desired. A bore 46 may pass through input gear 36 inorder to accommodate first half-shaft 28. It is contemplated that inputgear 36 could alternatively embody a spiral bevel gear, a straight bevelgear, a single- or double-helical gear, or a spur gear.

Carrier 38 may be a hollow and cylindrical housing that also functionsas a power-transmitting component. In particular, carrier 38 may be havea generally closed end 48 and an opposing end 50 that is open to aninternal cavity 52. An integral mounting flange 54 may be located atfirst end 48, and a plurality of fasteners 56 may be distributed aroundan outer periphery of carrier 38 to connect mounting flange 54 to inputgear 36. First and second side gears 40, 42, along with side pinions 44,may be assembled into cavity 52 via second end 50. A cap 58 may beconnected to second end 50 to enclose these components.

Each of first end 48 of carrier 38 and cap 58 may have a bore 60 formedtherein that is configured to receive first and second half-shafts 28,30. In some embodiments, a seal (not shown) may be installed withincarrier 38 and/or cap 58 around first and second half-shafts 28, 30 toinhibit debris ingress and lubrication leakage.

A plurality of cups 62 may be formed within an interior annular wall ofcarrier 38 (i.e., within a wall that annularly surrounds cavity 52). Inthe disclosed embodiment, carrier 38 includes four cups 62, and each cup62 is configured to receive a corresponding side pinion 44. It iscontemplated that a greater or lesser number of cups 62 could be formedwithin the interior annular wall of carrier 38, if desired. Cups 62 maybe co-located at about the same axial position, and equally distributedaround an interior periphery of carrier 38. In embodiments having fourcups 62, each cup 62 may be angularly space apart from each other byabout 90°. That is, each cup 62 may have a generally circular shape, andan axis 64 of each shape may be orthogonal to the axes 64 of adjacentshapes. In embodiments having a different number of cups 62, the spacingbetween axes 64 may be less than or more than 90°. For example, whenonly two cups 62 are included in carrier 38, axes 64 of cups 62 may beangularly spaced apart by about 180°. In the cross-sectional view ofFIG. 3, only three cups 62 are shown, and the corresponding side pinion44 is removed from the center cup 62 for clarity.

Each of cups 62 may have a generally flat bottom surface, and a curvedside wall that intersects with the flat bottom surface. In general, theshape of the curved side wall may conform to an external shape of sidepinion 44, such that a force directed diagonally through side pinion 44from side gears 40, 42 may be transmitted in a normal direction to thecurved side wall. In this way, a moment may not be created inside sidepinion 44 by the force. A clearance gap 66 having a substantiallyconstant thickness of about 0.0254 15 mm/0.001 in may be located betweenthe curved side wall of cup 62 and side pinion 44, and filled withlubrication during operation. An orifice 68 may be formed within theflat bottom surface of each cup and, as will be explained in more detailbelow, used as a conduit for lubrication. An inner diameter of orifice68 may be about equal to one-half of an inner diameter of the flatbottom surface of cup 62.

A plurality of additional orifices 70 may be formed in closed end 48 ofcarrier 38. In one embodiment, the number of additional orifices 70 maymatch the number of side pinions 44 installed in differential 22. Inanother embodiment, the number of additional orifices 70 may be somemultiple (e.g., 2×) of the number of side pinions 44. Orifices 70 shouldbe generally radially aligned with and/or extend to a mesh locationbetween the teeth of each pinion 44 and first side gear 40 (e.g.,aligned at least with an entrant location at which the gear teeth beginto mesh). As will be explained in more detail below, orifices 70 mayfacilitate lubrication of differential 22.

First and second side gears 40, 42 may be substantially identical bevelgears that are oriented in opposition to each other. External teeth ofboth gears 40, 42 may be configured to simultaneously intermesh with theteeth of all of side pinions 44 (e.g., at opposite sides of pinions 44)such that, as side pinions 44 are rotated with carrier 38 about axis 32,side gears 40, 42 may also be driven to rotate about the same axis 32. Abase end of first side gear 40 may be supported within a correspondingstep in carrier 38, while a base end of second side gear 42 may berotationally supported within a corresponding step in cap 58. Each offirst and second side gears 40, 42 may have a splined bore 72 formedtherein that is configured to receive a corresponding interior end offirst and second half-shafts 28, 30. With this configuration, a rotationof first and second side gears 40, 42 may result in a rotation of firstand second half-shafts 28, 30.

Side pinions 44 may be substantially identical bevel gears. As shown inFIGS. 4, 5, 6A, and 6B, each side pinion 44 may have a generally flatbottom 74, a generally flat top 76 located opposite bottom 74, aplurality of gear teeth 78 located adjacent top 76, and an arcuate outersurface (“surface”) 80 connecting bottom 74 with teeth 78. Surface 80may join gear teeth 78 at a transition region 82. An outer diameter attop 76 may be smaller than an outer diameter at bottom 74, and an outerdiameter at transition region 82 may be greater than the outer diameterat bottom 74.

Teeth 78 of side pinion 44 may have geometry that is at least partiallydefined by a plurality of angles. For example, a first angle α_(P) shownin FIG. 5 may represent a pressure angle of each tooth 78; a secondangle α_(F) shown in FIG. 6B may represent a face angle of teeth 78; athird angle α_(C) shown in FIGS. 5 and 6B may represent a cone angle ofteeth 78; and a fourth angle α_(R) shown in FIG. 6B may represent a rootangle of teeth 78. The pressure angle α_(P) may be defined as the anglebetween a primary pressure face of tooth 78 and a tangent to the pitchcircle of pinion 44. Each of the remaining angles may be defined by anarc extending between axis 64 and a corresponding line drawn through aPitch Apex (shown in FIGS. 5 and 6A) of side pinion 44 that lies on axis64.

As shown in FIGS. 4, 5, 6A, and 6B, when teeth 78 pinion 44 engage theteeth of side gears 40, 42, normal forces F_(N) (forces at the faces ofteeth 78 that are normal to the pressure angle α_(P)) may be generatedat opposing sides of pinion 44. These normal forces F_(N) may combine toproduce a resultant force F_(R). Resultant force F_(R) may lie in aplane (see FIG. 4) between the opposing sides of pinion 44, and may beinclined relative to bottom 74 and top 76 (see FIGS. 6A and 6B). If theresultant force F_(R) passes through surface 80 at an oblique anglerelative to a gradient at the surface, a moment could be created thatcauses tilting of pinion 44 when pinion 44 is pushed by the resultantforce F_(R) into cup 62.

The curvature of surface 80 may be designed such that resultant forceF_(R) passes through arcuate outer surface 80 at an angle that is alwaysnormal to surface 80. That is, arcuate outer surface 80 may be shapedsuch that resultant force F_(R) is oriented about 90° to surface 80 atthe application point of resultant force F_(R). In one embodiment,surface 80 may be at least partially defined by a curve 81 that isrotated about axis 64 of pinion 44. Curve 81 may be, for example, apolynomial curve of higher order (e.g., of 3^(rd) order or higher),defined by the following equation:p=Σ _(i=0) ^(n) a _(i) x ^(i)  Eq. 1

-   -   wherein:        -   p represents curve 81;        -   n represents the order of the curve;        -   a is a coefficient; and        -   x is distance along axis 64.

The angular orientation of resultant force F_(R) in space may be definedby an angle β (shown in FIG. 5) and controlled by the pressure angleα_(P) and the cone angle α_(C) according to the following equation:tan β=tan α_(P)·sin α_(C)  Eq. 2

-   -   wherein:        -   β is the angular orientation of resultant force F_(R) in            space;        -   α_(P) represents the pressure angle; and        -   α_(C) represents the pitch or cone angle.

Using Eq. 1 and Eq. 2 from above, curve 81 may be derived by requiringthat the cross product between a gradient (VS) of surface 80 and theresultant force F_(R) equals zero (i.e., that ∇S×F_(R)=0), at theapplication point of the resultant force F_(R).

It is contemplated that, in some applications, surface 80 may be dividedinto multiple sections that are axially adjacent each other. Forexample, surface 80 could be divided into a first section 80 a (see FIG.6A) located adjacent teeth 78, and a second section 80 b located betweenfirst section 80 a and bottom 74. In this example, section 80 b may belarger than section 80 a (e.g., 2-3 times larger). It should be notedthat surface 80 could be divided into more than two sections and/or thatthe sections could have different relative sizes, if desired. Each ofthe sections of surface 80 may have unique curvature defined by adifferent polynomial (e.g., a different order polynomial) and bedesigned to perform differently under different circumstances. Forexample, section 80 a may have a lower order polynomial than section 80b, as section 80 a may experience less loading the section 80 b. Ingeneral, a higher-power/lower-speed application may utilize surface 80created using more sections and higher-order polynomials (e.g., 5-6^(th)order), while a lower-power/higher-speed application may use fewersections (e.g., only one section) and lower-order polynomials (e.g.,3-4^(th) order). For the purpose of this application, ahigher-power/lower-speed application may be considered an applicationtransmitting about 447.4 kW/600 horsepower or more at speeds less thanabout 100 rpm.

In one specific exemplary embodiment of pinion 44, the pressure angleα_(P) may be about 14-25°; the face angle α_(F) may be about 35-37°; thecone angle α_(C) may be about 28-38°; and the root angle α_(R) may beabout 23-25°. It is contemplated that these angles may have differentvalues for different applications. For the purposes of this disclosure,the term “about”, when used in reference to a dimensional value, may bedefined as being within an acceptable range of manufacturing tolerances.

In the specific exemplary embodiment provided above, a radius r₁ may beused to at least partially define the curvature of surface 80 and mayhave its origin positioned along the trajectory of resultant forceF_(R). For example, the origin of radius r₁ may be positioned atcoordinates (dx, dy₁), which may be included in the vector of resultantforce F_(R), and located in the same plane as F_(R). In the disclosedembodiment, r₁ may be about 40-45 mm/1.575-1.771 in, while (dx, dy₁) maybe about equal to (15-20 mm/.591-.787 in, 90-95 mm/3.543-3.740 in) whenmeasured from axis 64 and the Pitch Apex of pinion 44, respectively. Inthis configuration, the origin of r₁ may be located a distance dy₂ frombottom 74 that is about 20-25 mm/.787-.984 in.

Axial ends of arcuate outer surface 80 may be designed to inhibit stressriser formation. For example, a lower edge connecting flat bottom 74with arcuate outer surface 80 may be rounded, and have a radius r₂ ofabout 1-3 mm/.039-.118 in. Likewise, transition region 82 that connectssurface 80 with teeth 78 may taper inward to have a back angle α_(B) ofabout 2-4°.

FIGS. 7 and 8 illustrate lubrication flows through differential 22during two different situations. Specifically, FIG. 7 illustrateslubrication flow during high-speed travel of machine 10, while FIG. 8illustrates lubrication flow during low-speed travel. During anyoperation of machine 10, differential 22 may be filled about halfway upwith lubrication. Specifically, the lubrication may fill a lower half ofan axle housing (not shown) in which differential 22 is located, and thelubrication may reach up to about axis 32 of carrier 38.

As shown in FIG. 7, during high-speed travel, when carrier 38 is drivento rotate at high rpms, the external surface of carrier 38 maycontinuously drag through the volume of lubrication inside the axlehousing. During this rotation, orifices 68 may fill with oil, and alayer of oil may cling to the outside surface of carrier 38. Thehigh-speed rotation of carrier 38 may result in high-speed rotations ofside pinions 44 and corresponding high-speed rotations of side gears 40,42. As the teeth of these gears mesh with each other, they may functionas pumps, drawing the lubrication from orifices 68 radially inward atseparation of the teeth and pushing the lubrication axially outward atengagement of the teeth. The lubrication may be discharged axiallyoutward through orifices 70. By this action, a lubrication film may becreated between side pinions 44 and cups 62 (i.e., within gaps 66), andalso between the teeth of the intermeshing gears.

When the travel of machine 10 transitions from high-speed travel tolow-speed travel, the direction of lubrication flow through differential22 may reverse. Specifically, when carrier 38 is driven to rotate at lowrpms, the teeth of input gear 36 may function as radially oriented fluidscrews (e.g., because of their pitch and helical angle). That is, theteeth of input gear 36 may be filled as they are dunked into the volumeof lubrication in the axle housing, and the pitch angle of the teeth,combined with the rotational motion of input gear 36, may overcome anycentrifugal forces acting on the lubrication and push the lubricationradially inward toward orifices 70. The lubrication may be pushedthrough the intermeshing engagement of side pinions 44 with side gears40, 42, and then be flung radially outward via through orifices 68. Inthis situation, the centrifugal forces flinging the lubrication radiallyoutward through orifices 68 may be greater than the pumping forces ofthe intermeshing teeth that would otherwise draw the lubricationradially inward. For the purposes of this disclosure, high-speed travelmay be considered travel at speeds within an upper two-thirds of anoverall speed range.

FIGS. 9 and 10 illustrate an alternative differential embodiment,labeled as element 84. Like differential 22 of FIGS. 1-8, differential84 of FIG. 9 may include input gear 36, first side gear 40, second sidegear 42, and side pinions 44. However, in contrast to differential 22,differential 84 may not include carrier 38. Instead, a nested carrier 86may be used to transfer torque from input gear 36 to side pinions 44. Inthis embodiment, nested carrier 86 may function primarily as apower-transmitting component, and additional housing members 88, 90 maybe used to enclose the other components of differential 84. Housingmembers 88, 90 may be substantially identical to each other (e.g.,within manufacturing tolerances), and bolted to opposing sides of inputgear 36 via fasteners 56. A plurality of orifices 92 may be formed inhousing members 88, 90 to function as conduits for lubrication. Becausehousing members 88, 90 may not be power-transmitting components, housingmembers 88, 90 may be thinner-walled and lighter weight, when comparedto carrier 38.

As shown in FIG. 9, differential 84 may be relatively compact. Inparticular, because carrier 86 may be nested radially inside of inputgear 36 (i.e., instead of being bolted at an axial end), the overallaxial length of differential 84 may be less. In this configuration,first side gear 40 may be located at a first axial end of input gear 36,while second side gear 42 may be located at a second axial end of inputgear 36 opposite the first axial end. In other words, first and secondside gears 40, 42 may be located at opposing sides of input gear 36,rather than both being located at the same side. This may afford greaterflexibility in packaging of drivetrain 16 (referring to FIG. 2). One ormore spacers 94 may be included to axially position nested carrier 86inside of input gear 36. Each of spacers 94 may be generally ring-like,and positioned between a corresponding one of housing members 88, 90 andan end of a corresponding one of side gears 40, 42.

For the sake of clarity, housing members 88 and 90, spacers 94, and sidegears 40 and 42 have been removed from FIG. 10. As seen in this figure,carrier 86 may be connected to input gear 36 by way of a splinedinterface 96. This type of interface may allow carrier 86 (as well asthe other components carried by carrier 86) to float axially somewhatwithin input gear 36 during assembly and operation of differential 84.This axial floating ability may account for manufacturinginconsistencies that could otherwise cause binding between carrier 86and input gear 36.

FIG. 11 illustrates an external side view of nested carrier 86, whileFIG. 12 illustrates a cross-sectional side view of nested carrier 86taken along a first axially oriented plane of symmetry 98. As shown inthese figures, nested carrier 86 may be generally cylindrical andhollow, having a two opposing open ends. In addition to being generallysymmetrical relative to plane 98, nested carrier 86 may also begenerally symmetrical relative to a second axially oriented plane 100and relative to a radially oriented plane 102. Nested carrier 86 mayhave formed therein a plurality of cups 104 (e.g., four cups 104) thatare identical to cups 62 described above in regard to differential 22.Specifically, each of cups 104 may include a flat bottom surface 106,and a curved side wall 108 that intersects with bottom surface 106. Ingeneral, the shape of curved side wall 108 may conform to an externalshape of side pinion 44, such that the reaction force F_(R) directeddiagonally through side pinion 44 from side gears 40, 42 may betransmitted in a normal direction to curved side wall 108. An orifice110 may be formed within bottom surface 106 of each cup 104 and, in amanner similar to what is explained above with respect to FIGS. 7 and 8,used for lubrication purposes. Outer edges of nested carrier 86 may berounded or beveled, as desired, to facilitate assembly into input gear36.

FIG. 13 illustrates another alternative differential embodiment, labeledas element 112. Like differential 84 of FIGS. 9 and 10, differential 112of FIG. 13 may include first side gear 40, second side gear 42 (omittedfrom FIG. 13 for the sake of clarity), and side pinions 44. However, incontrast to differential 84, differential 112 may include a differentinput gear 114, and a different nested carrier 116 used to transferrotation from input gear 114 to side pinions 44. Nested carrier 116,unlike nested carrier 86, may not form a complete cylinder. Instead,nested carrier 116 may be formed from two or more (e.g., from four)substantially identical carrier members 116 a that are annularly spacedapart from each other. Each carrier member 116 a may located inside apocket 118 formed within an interior of input gear 114, and a spacer 120may protrude radially inward to annularly separate adjacent carriermembers 116 a. In the disclosed embodiment, spacer 120 may be integralto input gear 114. In other embodiments, however, spacers 120 may beseparate standalone component.

As shown in FIG. 14, each carrier member 116 a may include externalsplines 122 configured to engage corresponding splines within pockets118 (referring to FIG. 13) of input gear 114, and a single cup 104. Eachcarrier member 116 a may be able to float axially somewhat relative toinput gear 114, yet still be connected to input gear 114 via splines 122to receive an input torque. By forming nested carrier 116 via multipleseparate members 116 a, the likelihood of misalignment or manufacturinginconsistencies causing binding of nested carrier 116 may be reducedeven further. It is contemplated that splines 122 could be omitted, ifdesired, and spacers 120 alternatively utilized for the torque transfer.

FIG. 15 illustrates another alternative differential embodiment, labeledas element 124. Like differential 84 of FIGS. 9 and 10, differential 124of FIG. 15 may include input gear 36, first side gear 40, second sidegear 42, side pinions 44, and housing member 90. However, in contrast todifferential 84, differential 124 may include a different nested carrier126 used to transfer rotation from input gear 36 to side pinions 44, alockup clutch 128 configured to selectively lock rotation of side gear42 to side gear 40 via nested carrier 126, and a larger (e.g., axiallylonger) housing member 130 used to enclose nested carrier 126 and lockupclutch 128.

Nested carrier 126, unlike nested carrier 86, may not be axiallysymmetric. In particular, nested carrier 126 may extend axially adistance in one direction past pinion gears 44 toward a base end of sidegear 42. One or more internal gear teeth (e.g., splines) 132 may beformed at an inner annular surface of the protruding portion of nestedcarrier 126. As will be explained in more detail below, teeth 132 may beused to selectively connect nested carrier 126 to second side gear 42via lockup clutch 128.

Lockup clutch 128 may itself be a sub-assembly of multiple componentsarranged to rotate around axis 32. For example, lockup clutch 128 mayinclude a disk stack 136, and a hydraulic actuator (not shown) that isconfigured to selectively compress disk stack 136. Disc stack 136 mayinclude a plurality of friction disks, a plurality of separator disksinterleaved with the friction disks and, in some instances, a damper(not shown) located at one or both ends of disc stack 136. The frictiondisks may be connected to rotate with one of nested carrier 126 (e.g.,via teeth 132) and second side gear 42 (e.g., by way of correspondingexternal teeth 134), while the separator disks may be connected torotate with the other of nested carrier 126 and second side gear 42. Inthis manner, when the hydraulic actuator is activated, the frictiondisks may be sandwiched between the separator disks, thereby creatingfriction that inhibits relative rotation between nested carrier 126 andsecond side gear 42. When the rotation of second side gear 42 isconstrained to the rotation of nested carrier 126, second side gear 42may rotate at the same speed as first side gear 40, regardless ofmachine turning or ground conditions. This may help to improve tractionduring poor ground conditions. A fluid pressure within the hydraulicactuator may relate to a magnitude of the friction resisting relativerotation, and be provide via one or more axial ports 138.

The hydraulic actuator may embody a service piston that works tocompress disk stack 136 under different conditions. The service pistonmay be ring-like and, together with housing member 130, form a controlchamber. When the control chamber is filled with pressurized oil viaport 138, the hydraulic actuator may be urged toward disk stack 136,thereby compressing disc stack 136.

In some embodiments, one or more springs (not shown) may be arranged invarious configurations to bias the hydraulic actuator away from discstack 136. In these configurations, when pressurized fluid is notsupplied into the control chamber, the hydraulic actuator may bedeactivated by the springs and moved away from disc stack 136 to reducethe friction generated between the disks and plates thereof.

Housing member 130, together with housing member 90, may substantiallyenclose the other rotating components of differential 124. As in thedifferential embodiment of FIGS. 9 and 10, housing members 90, 130associated with differential 124 may be used primarily to house theother components and not to transfer torque. For this reason, housingmembers 90, 130 may be relatively thin-walled and lightweight. Likehousing member 88, housing member 130 may additionally function toaxially position second side gear 42 by way of spacer 94.

INDUSTRIAL APPLICABILITY

The disclosed differentials and side pinions of the present disclosurehave potential application in any machine requiring torque delivery topaired traction devices. The disclosed differentials and side pinionsmay have extended life and improved efficiency. The extended life may beprovided by reducing, or even eliminating, moments within the sidepinions that tend to cock the side pinions away from desired positions.The moments may be eliminated by way of a unique outer curvature of theside pinions. The unique outer curvature of the side pinions may help toensure that the loading forces are directed normally into correspondingcups formed in the differentials. By maintaining the side pinions intheir desired positions, lubrication flow to the rotating components ofthe disclosed differentials may be ensured, which may prolong a life ofthese components and also reduce friction of the rotations. Reducedfriction may improve the efficiency of the disclosed differentials.

In addition, because the disclosed side pinions may not need spidershafts to support their rotations, the size, the weight, and the cost ofthe disclosed differentials may be lower. Further, without spidershafts, it may be easier to avoid moment creation and/or to distributethe loading forces over a greater area. This increased forcedistribution may reduce a likelihood of metallic contact betweencomponents, which may further reduce wear of the discloseddifferentials. Finally, because the disclosed side pinions may not havecentral bores normally required to accommodate spider shafts, alllubrication directed into the disclosed differentials at the sidepinions may be used to maintain an oil film between the side pinions andthe associated cups. This load-bearing layer of oil may further reducethe likelihood of metal-on-metal contact, which may prolong the life ofthe disclosed differentials.

Finally, the disclosed differentials may have greater packagingflexibility. Specifically, the axially compact design of some of thedisclosed differentials, may provide more space for other drivetraincomponent.

It will be apparent to those skilled in the art that variousmodifications and variations can be made to the differentials and sidepinions of the present disclosure without departing from the scope ofthe disclosure. Other embodiments will be apparent to those skilled inthe art from consideration of the specification and practice of thedifferentials and side pinions disclosed herein. It is intended that thespecification and examples be considered as exemplary only, with a truescope of the disclosure being indicated by the following claims andtheir equivalents.

What is claimed is:
 1. A side pinion for a differential, comprising: abody having a flat bottom and a flat top located at an end opposite theflat bottom; a plurality of gear teeth formed adjacent the flat top; andan arcuate outer surface connecting the plurality of gear teeth to theflat bottom; wherein the arcuate outer surface is formed by rotating apolynomial curve of third order or higher around an axis of the sidepinion.
 2. The side pinion of claim 1, wherein: the differential hasside gears that are configured to engage the plurality of gear teeth atopposing sides of the side pinion; when the side gears engage theplurality of gear teeth at the opposing sides of the side pinion, areaction force is created; and the polynomial curve is selected suchthat a gradient of the arcuate outer surface at an application point ofthe reaction force is about aligned with the reaction force.
 3. The sidepinion of claim 2, wherein: a pressure angle of the plurality of gearteeth is about 14-25°; a root angle of the plurality of gear teeth isabout 23-25°; and a cone angle of the side pinion is about 28-38°. 4.The side pinion of claim 3, wherein a radius of the arcuate outersurface at an application point of the reaction force is about 40-45 mm.5. The side pinion of claim 4, wherein an edge radius at a transition ofthe arcuate outer surface and the flat bottom is about 1-3 mm.
 6. Theside pinion of claim 5, wherein: the arcuate outer surface joins theplurality of gear teeth at a transition region; and the transitionregion tapers inward away from the arcuate outer surface at a back angleof about 2-4°.
 7. The side pinion of claim 1, wherein the arcuate outersurface comprises multiple sections that are located axially adjacent toeach other and that are each defined by a different polynomial curve. 8.The side pinion of claim 7, wherein a first section of the arcuate outersurface adjacent the plurality of gear teeth is defined by a polynomialcurve having a lower order than a second section of the arcuate outersurface located between the first section and the bottom.
 9. The sidepinion of claim 1, wherein: the differential is designed for usein-high-power applications of 447.4 kW/600 horsepower or higher; and thearcuate outer surface comprises multiple sections formed by a 5^(th) or6^(th) order polynomial curve.
 10. The side pinion of claim 1, wherein:the differential is designed for use in-low-power applications of 447.4kW/600 horsepower or less; and the arcuate outer surface comprises asingle section formed by a 3^(rd) or 4^(th) order polynomial curve. 11.The side pinion of claim 1, wherein an outer diameter at the flat top issmaller than an outer diameter at the flat bottom.
 12. The side pinionof claim 11, wherein an outer diameter at an axial transition regionbetween the plurality of gear teeth and the arcuate outer surface islarger than the outer diameter at the flat bottom.
 13. A differential,comprising: an input gear; a carrier fixedly connected to the input gearand configured to rotate together with the input gear about a primaryaxis, the carrier having an internal annular surface with a plurality ofcups formed therein and equally spaced apart around a circumference ofthe carrier; a first side gear disposed inside the carrier andconfigured to rotate about the primary axis; a second side gear disposedin the carrier at an end opposite the first side gear and alsoconfigured to rotate about the primary axis; a side pinion having a bodyhaving a flat bottom and a flat top located at an end opposite the flatbottom, the side pinion disposed within each of the plurality of cupsand intermeshed with both of the first and second side gears, the sidepinion having a plurality of gear teeth that protrude from an associatedone of the plurality of cups radially inward toward the primary axis,and an arcuate outer surface connected to the plurality of gear teeth atan axial transition region and conforming to an inner contour of theplurality of cups; wherein the arcuate outer surface is formed byrotating a polynomial curve of third order or higher around an axis ofthe side pinion.
 14. The differential of claim 13, wherein: when thefirst and second side gears engage the plurality of gear teeth atopposing sides of the side pinion, a reaction force is created; and thepolynomial curve is selected such that a gradient of the arcuate outersurface at an application point of the reaction force is about alignedwith the reaction force.
 15. The differential of claim 13, wherein thearcuate outer surface of the side pinion comprises multiple sectionsthat are located axially adjacent to each other and that are eachdefined by a different polynomial curve.
 16. The differential of claim15, wherein a first section of the arcuate outer surface adjacent theplurality of gear teeth is defined by a polynomial curve having a lowerorder than a second section of the arcuate outer surface located betweenthe first section and the bottom.
 17. The differential of claim 13,wherein: an outer diameter at the flat top is smaller than an outerdiameter at the flat bottom; and an outer diameter at the axialtransition region is larger than the outer diameter at the flat bottom.18. A drivetrain for a mobile machine having first and second tractiondevices located at opposing sides, the drivetrain comprising: a powersource; a transmission driven by the power source; a main pinionoperatively connected to an output of the transmission; a firsthalf-shaft connected to the first traction device; a second half-shaftconnected to the second traction device; and a differential driven bythe main pinion to rotate the first and second half-shafts withsubstantially equal torque, the differential including: an input gearintermeshed with the main pinion; a carrier fixedly connected to theinput gear and configured to rotate together with the input gear about aprimary axis, the carrier having an internal annular surface with aplurality of cups formed therein and equally spaced apart around acircumference of the carrier; a first side gear disposed inside thecarrier and having external teeth at an inner end and an outer endconnected to the first half-shaft; a second side gear disposed insidethe carrier and having external teeth at an inner end and an outer endconnected to the second half-shaft; and a side pinion disposed withineach of the plurality of cups and intermeshed with the external teeth ofboth of the first and second side gears, the side pinion having aplurality of gear teeth that protrude from an associated one of theplurality of cups radially inward toward the primary axis, and anarcuate outer surface connected to the plurality of gear teeth at anaxial transition region and conforming to an inner contour of theplurality of cups, wherein: the arcuate outer surface of the side pinionis formed by rotating a polynomial curve of third order or higher aroundan axis of the side pinion; when the first and second side gears engagethe plurality of gear teeth at opposing sides of the side pinion, areaction force is created; and the polynomial curve is selected suchthat a gradient of the arcuate outer surface at an application point ofthe reaction force is about aligned with the reaction force.